A roller cone rock bit is a commonly used cutting tool used in oil, gas, and mining fields for breaking through earth formations and shaping well bores. Reference is made to FIG. 1 which illustrates a cross-sectional view of a portion of a typical roller cone rock bit. FIG. 1 specifically illustrates the portion comprising one head and cone assembly of the bit. The general configuration and operation of such a bit is well known to those skilled in the art.
The head 10 of the bit includes a downwardly and inwardly extending bearing shaft 12. A cutting cone 14 is rotatably mounted on the bearing shaft 12. The bearing system for the head and cone assembly that is used in roller cone rock bits to rotatably support the cone 14 on the bearing shaft 12 typically employs either rollers as the load carrying element (a roller bearing system) or a journal as the load carrying element (a friction bearing system). FIG. 1 specifically illustrates a friction journal bearing implementation including a bearing system defined by a first cylindrical friction bearing 16 (also referred to as the main journal bearing). The cone 14 is axially retained on the bearing shaft 12, and further supported for rotation, by a set of ball bearings 18 provided within an annular raceway 20. The bearing system for the head and cone assembly further includes second cylindrical friction bearing 22, first radial friction (thrust) bearing 24 and second radial friction (thrust) bearing 26.
The bearing system for the head and cone assembly of the bit is lubricated and sealed. The interstitial volume within the bearing system defined between the cone 14 and the bearing shaft 12 is filled with a lubricant (typically, grease). This lubricant is provided to the interstitial volume through a series of lubricant channels 28. A pressure compensator 30, usually including an elastomer diaphragm, is coupled in fluid communication with the series of lubricant channels 28. The lubricant is retained within the bearing system by a sealing system 32 provided between the base of the cone 14 and the base of the bearing shaft 12. The configuration and operation of the lubrication and sealing systems within roller cone drill bits are well known to those skilled in the art.
A body portion 34 of the bit, from which the head and cone assembly depends, includes an upper threaded portion forming a tool joint connection which facilitates connection of the bit to a drill string (not shown, but well understood by those skilled in the art).
FIG. 2 illustrates a cross-sectional view of the bit shown in FIG. 1 focusing on a portion of the bearing system in greater detail. In particular, FIG. 2 specifically focuses on the area of the first cylindrical friction bearing (main journal bearing) 16. The first cylindrical friction bearing 16 is defined by an outer cylindrical surface 40 on the bearing shaft 12 and an inner cylindrical surface 42 of a bushing 44 which has been press fit into the cone 14. This bushing 44 is a ring-shaped structure typically made of beryllium copper, although the use of other materials is known in the art. In a roller bearing system, the outer cylindrical surface 40 on the bearing shaft 12 would interact with roller bearings maintained, for example, in an annular roller raceway within the cone 14.
FIG. 2 further shows that the ball bearings 18 ride in the annular raceway 20 defined at an interface between the bearing shaft 12 and cone 14. The ball bearings 18 are delivered to the raceway 20 through a ball opening 46, with that opening 46 being closed by a ball plug 48. The ball plug 48 is shaped to define a portion of the lubricant channels 28 within the ball opening 46. The ball bearing system as shown would typically also present in bearing system implementations which utilize roller bearings.
As discussed above, lubricant is retained within the bearing system by a sealing system 32. The sealing system 32, in a basic configuration, comprises an o-ring type seal member 50 positioned in a seal gland 52 between the cutter cone 14 and the bearing shaft 12 to retain lubricant and exclude external debris. A cylindrical surface seal boss 54 is provided at the base of the bearing shaft 12. In the illustrated configuration, this surface of the seal boss 54 is outwardly radially offset (for example, by the thickness of the bushing 44) from the outer cylindrical surface 40 of the first friction bearing 16. It will be understood that the seal boss 54 could exhibit no offset with respect to the main journal bearing 16 surface 40 if desired. The annular seal gland 52 is formed in the base of the cone 14. The gland 52 and seal boss 54 align with each other when the cutting cone 14 is rotatably positioned on the bearing shaft 12. The o-ring sealing member 50 is compressed between the surface(s) of the gland 52 and the seal boss 54, and functions to retain lubricant within the bearing system. This sealing member 50 also prevents materials in the well bore (such as drilling mud and debris) from entering into the bearing system.
Over time, the rock bit industry has moved from a standard nitrile material for the seal member 50, to a highly saturated nitrile elastomer for added stability of properties (thermal resistance, chemical resistance). The use of a sealing system 32 in rock bit bearings has dramatically increased bearing life in the past fifty years. The longer the sealing system 32 functions to retain lubricant within the interstitial volume, and exclude contamination of the bearing system, the longer the life of the bearing and drill bit. The sealing system 32 is, thus, a critical component of the rock bit.
With reference once again to FIG. 1, the second cylindrical friction bearing 22 of the bearing system is defined by an outer cylindrical surface 60 on the bearing shaft 12 and an inner cylindrical surface 62 on the cone 14. The outer cylindrical surface 60 is inwardly radially offset from the outer cylindrical surface 40 (FIG. 2). The first radial friction bearing 24 of the bearing system is defined between the first and second cylindrical friction bearings 16 and 22 by a first radial surface 64 on the bearing shaft 12 and a second radial surface 66 on the cone 14. The second radial friction bearing 26 of the bearing system is adjacent the second cylindrical friction bearing 22 at the axis of rotation for the cone and is defined by a third radial surface 68 on the bearing shaft 12 and a fourth radial surface 70 on the cone 14.
The lubricant is provided in the interstitial volume that is defined between the surfaces 40 and 42 of the first cylindrical friction bearing 16, the surfaces 60 and 62 of the second cylindrical friction bearing 22, the surfaces 64 and 64 of the first radial friction bearing 24 and the surfaces 68 and 70 of the second radial friction bearing 26. The sealing system 32 with the o-ring type seal member 50 positioned in the seal gland 52 functions to retain the lubricant within the lubrication system and specifically between the opposed radial and cylindrical surfaces of the bearing system.
During operation of the bit, the rotating cone 14 oscillates along the head in at least an axial manner. This motion is commonly referred to in the art as a “cone pump.” Cone pumping is an inherent motion resulting from the external force that is imposed on the cone by the rocks during the drilling process. The oscillating frequency of this cone pump motion with respect to the head is related to the rotating speed of the bit. The magnitude of the oscillating cone pump motion is related to the manufacturing clearances provided within the bearing system (more specifically, the manufacturing clearances between the surfaces 40 and 42 of the first cylindrical friction bearing 16, the surfaces 60 and 62 of the second cylindrical friction bearing 22, the surfaces 64 and 64 of the first radial friction bearing 24 and the surfaces 68 and 70 of the second radial friction bearing 26). The magnitude is further influenced by the geometry and tolerances associated with the retaining system for the cone (for example, the ball race). When cone pump motion occurs, the interstitial volume defined between the foregoing cylindrical and radial surfaces of the bearing system changes. This change in volume squeezes the lubricant provided within the interstitial volume. The change in interstitial volume and squeezing of the lubricant grease results in the generation of a lubricant pressure pulse. Over a very short period of time, responsive to this pressure pulse, grease flows along a first path between the bearing system and the pressure compensator 30 through the series of lubricant channels 28. The pressure compensator 30 is designed to relieve or dampen the pressure pulse by compensating for volume changes through its elastomer diaphragm. However, it is known in the art that the pressure pulse, notwithstanding the presence and actuation of the pressure compensator 30, can also be felt at the sealing system 32 due to the presence of a separate second path for the flow of grease, responsive to this pressure pulse, between the opposed radial and cylindrical surfaces of the bearing system and the sealing system 32.
The flow of grease along this second path in response to the pressure pulse is known to be detrimental to seal operation and can also reduce seal life. For example, positive and negative pressure pulses due to cone pump motion may cause movement of the sealing member 50 within the seal gland. A nibbling and wearing of the seal member 50 may result from this movement. Additionally, a positive pressure pulse due to cone pump motion may cause lubricant grease to leak out past the sealing system 32. A negative pressure pulse due to cone pump motion may pull materials from the well bore (such as drilling mud and debris) past the sealing system 32 and into the bearing system.
Reference is now made to FIG. 3 which shows a cross-section of the bearing shaft 12 generally at the location of the first friction bearing 16 taken along dotted line 80 of FIG. 2. As is known by those skilled in the art, the first friction bearing 16 for the bearing system includes a loading zone (having an arc angle of about 120°-180°) which bears the load of the cone 14 and a non-loading zone (having an arc angle of about 180°-240°). The outer surface 40 of the bearing shaft 12 at the loading zone is typically hardfaced (not explicitly shown, but known to those skilled in the art). One of the lubricant channels 28 for the lubrication system terminates at the outer cylindrical surface 40 of the bearing shaft 12 in the area of the non-loading zone. The termination of the lubricant channel 28 on the outer surface 40 of the bearing shaft 12 is typically provided by a circumferentially positioned groove 90 that is milled or machined into the outer surface 40. This groove 90 includes an opening 92 for providing fluid communication into the lubricant channel 28.
Reference is now made to FIG. 4 which shows a side view of the bearing shaft 12 focusing on the non-loading zone. The circumferentially positioned groove 90 terminates the lubricant channel 28 at the outer surface 40 of the first friction bearing 16 for the bearing system using opening 92. The axial width 94 of the groove 90 spans most, but not all, of the axial width 96 of the surface 40 for the first friction bearing 16 of the bearing system. For example, the axial width 94 is typically equal to the axial width 96 minus a constant (such as twice a fraction of an inch, for example, 2* 1/32″ or 2* 3/64″. In this way, the axial width 94 is typically greater than 80-90% of the axial width 96. The groove 90 is typically axially centered with respect to the surface 40 providing two equally sized attenuation zones 100. Because of the relative widths 94 and 96, the attenuation zones 100 present a minimal amount of outer surface 40 for the first friction bearing 16 that is located axially adjacent the groove 90 and present along the path shown by arrow 98. This minimal amount of outer surface 40 is insufficient to restrict the flow of grease and the passage of a pressure pulse between the bearing system (at surfaces 60, 64 and 68) and the sealing system 32 (at surface 54) along path 98. More specifically, this minimal amount of surface 40 along the path of arrow 98 provides only two relatively short (in an axial direction) attenuation zones 100 which might assist in attenuating the flow of grease along the path of arrow 98 resulting from the axial passage of the pressure pulse. In this configuration, the pressure pulse may travel along surface 40 and reach the sealing system 32 (at surface 54) before being dampened by the pressure compensator 30. As discussed above, this pressure pulse may have detrimental effects on the sealing system 32 and particularly the sealing member 50. There is accordingly a need in the art to reduce, or eliminate, the pressure pulsation due to cone pumping from acting on the sealing system 32.